Heat engine with external hot source

ABSTRACT

A heat engine with external hot source in which the engine has at least one variable volume working chamber for a working gas, and a distribution mechanism that connects this chamber to a cold input from an energy receiving path during an outgoing transfer phase and to a hot output of the energy receiving path during an incoming transfer phase, the energy receiving path being intended to heat the working gas outside the chamber on contact with the external hot source, wherein the distribution mechanism is timed in such a way as to: maintain pressure in the energy receiving path; during stable operation, connect the working chamber with the cold input from the energy receiving path whilst the pressure in the chamber is lower than the pressure in the energy receiving path.

The present invention relates to a heat engine with external hot source,in particular an exhaust heat recovery hot source, for applications onall types of vehicle, whether terrestrial, marine or airborne.

The operation of conventional hot source heat engines is above alleffective when the hot source is at a high temperature. Theirperformance declines however when the temperature of said hot source ismoderate, as is the case of the exhaust gases from internal combustionengines.

Internal combustion engines have relatively modest energy output. Thisis largely due to the thermal energy that these engines release into theenvironment through their cooling systems and especially through theirexhausts.

Particularly for vehicle applications, in which the weight, bulk andcost of the engine play an important part, the energy released has untilnow been under-exploited. For example, the energy released at theexhaust is in the form of a low pressure gas with a relatively moderatetemperature. The conversion of this released energy into mechanicalenergy by normal means involves heavy, bulky, costly machinery ofdebatable efficiency.

Solutions such as those described by U.S. Pat. No. 3,180,078 A, U.S.Pat. No. 4,121,423 A, DE 101 43 342 A1, JP 2004 270625 A, and U.S. Pat.No. 4,754,606 A do however propose hybrid heat-internal combustionengine solutions. However, none of these solutions is capable ofovercoming all of these weight, bulk, cost and above all, efficiencyconstraints. Most of this prior art envisages compressing a working gasand then reheating it with recovered heat before expanding it. However,as the recovered heat is available at a temperature that is onlyslightly higher than the temperature of the working gas at the end ofcompression, this is an inefficient process.

The object of this invention is thus to propose a heat engine capable ofefficiently converting thermal energy originating from a warm source,typically the waste heat released at the end of the internal combustionprocess, into mechanical energy, in particular in a manner compatiblewith the normal requirements of applications to vehicles.

According to the invention, the heat engine with external hot source inwhich the engine has at least one variable volume working chamber for aworking gas, and a distribution mechanism that connects this chamber toa cold input from an energy receiving path during an outgoing transferphase and to a hot output of the energy receiving path during anincoming transfer phase, the energy receiving path being intended toheat the working gas outside the chamber on contact with the externalhot source, is characterised in that the distribution mechanism is timedin such a way as to:

-   -   maintain pressure in the energy receiving path;    -   during stable operation, connect the working chamber with the        cold input of the energy receiving path whilst the pressure in        the chamber is lower than the pressure in the energy receiving        path.

The valve timing according to the invention is surprising as it allowsworking gas already located in the exchange path to flow back into theworking chamber, following which the working gas that has flowed backand the working gas that has just been compressed must both be expelledin the exchange path. Consequently, at first glance this results in anincrease in the negative work of the cycle. However, the working gas,which was slightly compressed before the chamber was connected to thecold input of the exchange path, is therefore at a relatively lowtemperature. Due to its low temperature, the working gas that has justbeen (relatively slightly) compressed is able to collect a relativelylarge quantity of thermal energy from the exhaust gases of the internalcombustion engine, despite their relatively low temperature. It istherefore possible to almost multiply by two the absolute temperature ofthe working gas in the exchange path, and therefore multiply by two thevolume of gas at the beginning of expansion relative to the volume ofgas at the end of compression, for the same mass of working gas. Theinvention thus allows for a thermodynamic cycle to be produced in theheat engine with a relatively large area, therefore producingsignificant mechanical power.

Furthermore, the working gas at the end of expansion can be of a highertemperature than the working gas at the end of compression and at thebeginning of the exchange path. This phenomenon allows for the workinggas discharged by the heat engine piston to be used as a heat source fora first stage of the external heat source, in parallel or combined withthe exhaust gases from the heat engine.

Due to this simple, low-cost solution, the heat engine is capable ofimproved nominal performance whilst allowing for lower weight andsmaller bulk. It is therefore completely compatible with the normaldemands of vehicle applications, and can be combined with the vehicle'sinternal combustion engine to form a hybrid engine.

This combination can be achieved in a preferred architecture of thehybrid engine in which the internal combustion engine comprises pistonscoupled to a shaft of the hybrid engine, and the heat engine has atleast one piston coupled to the shaft of the hybrid engine. The hotsource of the heat engine is then supplied with heat energy by theexhaust of the internal combustion engine.

Due to this architecture, the general structure of a hybrid enginediffers little from that of a conventional internal combustion engine,for example of a rod and crank type.

The bulk, weight and cost of the whole unit are compatible with currentrequirements, whilst the specific fuel consumption of the hybrid engine(quantity of fuel consumed per unit of power and unit of time) isparticularly low.

In particular, the internal combustion engine and the heat engine canhave a common engine block in which are formed cylinders of identicalsize, in which the pistons have strokes of identical length. It can alsobe envisaged that the diameter of the cylinders of the heat engine,and/or the stroke of its pistons, differ slightly from those of theinternal combustion engine, even if the heat engine and the internalcombustion engine have a common engine block.

For example, an engine with three cylinders can be envisaged, in whichtwo cylinders belong to the internal combustion engine and one cylinderbelongs to the heat engine, recovering and exploiting the exhaust energyfrom the two cylinders of the internal combustion engine.

Generally, the heat engine can operate with a suitable mass of workinggas to absorb the heat energy that can be recovered in the exhaust ofthe internal combustion engine. This mass of working gas can becontrolled by appropriate supercharging and/or by the selection of atwo-stroke cycle, even if the internal combustion engine operates on afour-stroke cycle, and/or by a specific swept volume (displacement) forthe heat engine.

In a preferred version, the heat engine comprises:

-   -   a working chamber delimited by the piston on the heat engine,        which alternately causes chamber volume growth and reduction        strokes;    -   a heat exchange path to heat a working gas outside the chamber,        on contact with the external hot source;    -   a distribution mechanism to selectively close the working        chamber and respectively selectively connect the working chamber        with a working gas intake, a working gas discharge, a cold input        of the exchange path and a hot output of the exchange path.

The exhaust gases from the internal combustion engine and/or the workinggas discharged by the heat engine then preferably enter the turbine of aturbocharger. The compressor of the turbocharger supplies the intake ofthe heat engine and/or the intake of the internal combustion engine.

The internal combustion engine typically operates on an Otto cycle or aDiesel cycle.

It is advantageous that the valve timing be controllable, particularlyto allow for initial pressurisation of the exchange path. To this end,the mass of gas sent to the exchange path on outgoing transfer must begreater than the mass of gas taken from the exchange path on incomingtransfer until the exchange path reaches the desired pressure. The valvetiming can also form part of a control system, for example a pressurecontroller in the exchange path.

The heat engine can operate on a two-stroke cycle in which the dischargephase, the compression phase and the outgoing transfer phase follow onfrom one another during a single chamber volume reduction stroke. Theintake into the working chamber can then take place at relatively highpressure between the discharge phase and the compression phase, throughsupercharging producing relatively high pressure.

The heat engine can also operate on a four-stroke cycle. In this case,instead of corresponding to a brief phase between discharge andcompression, the working gas intake can occupy a complete chamber volumegrowth stroke. During the next stroke, the reduction in the volume ofthe chamber, the compression phase is followed by the outgoing transferphase.

During the next stroke, the incoming transfer phase is followed by theexpansion phase, whilst the fourth stroke corresponds to the dischargeof the working gas from the chamber.

Other features and advantages of the invention will become apparent fromthe following description, which relates to non-limitative examples.

In the attached drawings:

FIG. 1 is a schematic diagram of a hybrid engine according to theinvention, with a 1 cylinder, two-stroke heat engine;

FIG. 2 is a diagrammatic longitudinal cross-sectional view of the enginein FIG. 1;

FIG. 3 is an end view of the crankshaft of the engine in FIGS. 1 and 2;

FIG. 4 is a diagram showing a two-stroke cycle of the heat engine of ahybrid engine according to the invention;

FIG. 5 is a diagrammatic axial view of the cylinder of the heat engine;

FIG. 5A is a vertical cross-sectional view of the heat engine of ahybrid engine according to the invention, along the plane A-A in FIG. 5,passing through the intake and discharge ports, during a dischargephase;

FIG. 5B is a similar view to that in FIG. 5A but the verticalcross-section is along the plane B-B in FIG. 5, passing through theoutgoing transfer and incoming transfer ports of the heat engine, stillduring the discharge phase of the cycle;

FIGS. 6A to 10B are all longitudinal cross-sections of the heat engine,along the plane A-A in FIG. 5 when the figure has a number followed bythe letter A, and respectively along the plane B-B in FIG. 5 when thefigure has a number followed by the letter B, in the following cyclephases:

FIGS. 6A and 6B: intake and scavenging phase;

FIGS. 7A and 7B: compression phase;

FIGS. 8A and 8B: outgoing transfer phase;

FIGS. 9A and 8B: incoming transfer phase;

FIGS. 10A and 10B: expansion phase;

FIG. 11 is a similar diagram to FIG. 4, but showing a four-stroke cycleof the heat engine according to the invention;

FIG. 12 is a partial vertical cross-sectional view of the top of theinternal combustion engine according to the invention;

FIG. 13 is a cross-sectional view along the line XIII-XIII in FIG. 12;and

FIGS. 14 to 17 are diagrams of four possible configurations of hybridengines according to the invention.

The example of a hybrid engine according to the invention shown in FIGS.1 to 3 in an extremely diagrammatic manner comprises a common engineblock 1 in which three parallel cylindrical bores, 2 c, 2 t are formedin a so-called “in-line” arrangement, that is, the axes of the bores areco-planar. According to the normal terminology, the cylindrical bores 2c, 2 t are known as “cylinders”. In its lower region, the engine blockhas bearings 3 aligned along an axis 7, which hold a crankshaft 4 thatis common to all of the cylinders 2 c, 2 t. The crankshaft 4 comprises ajournal 6 facing each of the cylinders 2 c, 2 t. As shown in FIG. 3, inthis three-cylinder example, the journals 6 are distributed at an angleof 120° to each other, that is evenly, around the axis of rotation 7 ofthe crankshaft.

According to the invention, the hybrid engine brings together aninternal combustion engine to which, in this example, the two cylinders2 c located at the ends are allocated, and a so-called “heat” engineaccording to the invention, to which the central cylinder 2 t isallocated.

A sliding piston 8 c is mounted in each of the cylinders 2 c of theinternal combustion engine. A sliding piston 8 t is mounted in thecylinder 2 t of the heat engine. Each piston 8 c, 8 t is connected tothe corresponding journal 6 by a connecting rod 9 coupled on the onehand to the piston and on the other hand to the journal.

In the specific example described, the cylinders 2 c and 2 t haveidentical bore diameters and their working strokes are identical lengths(equal to twice the radius of eccentricity of the journals 6). Theytherefore have equal displacements (volume swept by each piston in itscylinder). Preferably, provision is made for the connecting rods 9 to beidentical and the pistons 8 c and 8 t to have equal mass, so that thehybrid engine does not pose any particular dynamic balancing problems.

The engine block 1 is topped by a cylinder head 11 that is only verydiagrammatically shown in FIG. 2. The cylinder head 11 comprises intakepassages 12 and exhaust passages 13 for each of the cylinders 2 c of theinternal combustion engine, together with an intake passage 14, adischarge passage 16, an outgoing transfer passage 17 and an incomingtransfer passage 18 for the cylinder 2 t of the heat engine.

The internal combustion engine typically operates on a conventional Ottoor Diesel cycle and, in the example shown, each cylinder of the internalcombustion engine is associated with two intake ports 19 and two exhaustports 21, each fitted with an intake or exhaust valve 20 respectively.

Each of the passages 14, 16, 17, 18 of the heat engine is associatedwith a port 24, 26, 27, 28 that opens the corresponding passage into theworking chamber 22 of the heat engine, defined between the pressure face23 of the piston 8 t, the wall of the bore 2 t and the lower face of thecylinder head 11. Each port 24, 26, 27, 28 is fitted with a valve 29.When the four valves 29 (only two are shown in FIG. 2) are closed, thatis close the ports with which they are respectively associated, theworking chamber 22 is hermetically sealed.

In the example shown, the valves 20 and 29 are as a whole aligned in tworows so that they are controlled by two camshafts 31, one of which isshown in FIG. 2, and the position of which is simply represented inFIG. 1. By conventional means, the camshafts 31 are coupled to thecrankshaft 4 so that they rotate half as quickly as the crankshaft 4 inthis example, in which the internal combustion engine operates on afour-stroke cycle.

The heat engine, corresponding to the central cylinder 2 t, isassociated with an external hot source 32 (FIG. 1) made up of a gas-gasheat exchanger comprising an energy receiving path 33 along whichtravels gas, generally air, that has been previously compressed in theworking chamber 22 of the heat engine, and an energy supply path 34along which gases travel which are mainly the exhaust gases from theinternal combustion engine, therefore originating directly or indirectlyfrom the exhaust passages 13 and the exhaust ports 21 associated withthe bores 2 c of the internal combustion engine. The exchanger formingthe hot source 32 is a counterflow exchanger, that is, the gasoriginating from the bore 2 t of the heat engine passes through theexchanger in the opposite direction to the exhaust gas from the internalcombustion engine, as shown by the arrows in FIG. 1. The input of theenergy receiving path 33, connected to the outgoing transfer port 27, isin thermal contact with the cold output of the path 34. The output ofthe energy receiving path 33, connected to the incoming transfer port28, is in thermal contact with the hot input of the energy supply path34.

In the example shown more particularly in FIG. 1, the internalcombustion engine takes in atmospheric air and its exhaust is connectedto the turbine part 36 of a turbocharger 37. At the output of theturbine 36 the exhaust gases from the internal combustion engine passinto a catalytic converter 41 before reaching the energy supply path 34of the hot source 32. At the output of the hot source 32, the exhaustgases from the internal combustion engine are released into theatmosphere as shown by the arrow 42, generally through a silencersystem, not shown.

The compressor part 38 of the turbocharger 37 takes in air from theatmosphere and sends compressed air to the intake port 24 of the heatengine, by means of a charge-air cooler (intercooler) 43.

The thermodynamic cycle of the heat engine in the example of atwo-stroke engine will now be described with reference to FIG. 4.

FIG. 4 shows on the vertical axis the pressure P in the working chamber22 and on the horizontal axis the volume V of the working chamber asdefined in a variable manner by the piston 8 t.

It is a two-stroke cycle as all of the phases of the cycle take place ina single revolution of the crankshaft 4 and consequently in a singlereciprocation of the piston 8 t. The direction of travel of the cycle isindicated by arrows on the closed curve illustrating the cycle in FIG.4. Starting from a position in which the working chamber 22 is at itsmaximum volume VM (position shown in FIG. 2 with the piston 8 t as faraway as possible from the cylinder head 11), the first phase P1 of thecycle is a discharge phase, illustrated in FIGS. 5A and 5B, in which thedischarge port 26 is opened by the corresponding valve 29, whilst theother ports 24, 27, 28 are closed by their respective valves 29. The gas(air) contained in the working chamber 22 is discharged into thedischarge passage 16 by the rising of the piston 8 t. As shown in FIG.1, the passage 16 can connect to the exhaust outlet 42 of the internalcombustion engine. In a preferred variant, the passage 16 is connectedto a branch 44 that opens into the energy supply path 34 of the hotsource 32. The temperature of the exhaust gases decreases graduallyalong the path 34, from its input to its output. The branch 44 is a“warm” input positioned in an area where the temperature of the exhaustgases is between the temperature of the input and output of the path 34,and approximately equal to the temperature at which the gas isdischarged from the heat engine.

More specifically, the temperature at the input of the energy supplypath 34 can be 800° C., and it can be in the region of 200° C. at theoutput of that path. Along the path 34, the temperature of the exhaustgases therefore gradually decreases from 800° C. to 200° C. If the gasesdischarged by the heat engine have a temperature in the region of 300°C., the branch 44 is opened out into the area of the path 34 where theexhaust gases have a temperature of 300° C.

Given that the camshafts 31 rotate half as quickly as the crankshaft 4of the engine whilst the heat engine operates in this example of atwo-stroke cycle corresponding to a single revolution of the crankshaft,a complete cycle of the heat engine corresponds to a half-revolution ofthe camshafts. This is why, as shown in FIGS. 5A and 5B, the cams 44held by the camshafts 31 and associated with the heat engine are of atype with two diametrically opposed profiles to perform two operatingcycles per revolution.

The discharge phase P1 is followed by an intake and scavenging phase P2in which, as shown in FIGS. 6A and 6B, the intake 24 and discharge 26ports are simultaneously opened by their respective valves 29 so thatthe air produced by the compressor part 38 of the turbocharger 37 entersthe working chamber 22 and pushes the remaining exhaust gases out of it.During this time, the two transfer ports 27 and 28 are still closed bytheir corresponding valves 29.

Phase P2 is followed by an adiabatic compression phase P3 (FIGS. 7A and7B) in which the piston 8 t continues its stroke towards the cylinderhead 11, reducing the volume of the working chamber 22. During thiscompression phase, the four ports 24, 26, 27 and 28 are closed by theirrespective valves 29 and the working chamber 22 is hermetically sealed.

When the working chamber 22 reaches a volume Vs greater than its minimumvolume Vm, the outgoing transfer port 27 (FIG. 8B) is opened by thecorresponding valve 29, whilst the other ports remain closed. Thisresults in a back-flow phase P4 during which pressurised gas containedin the energy receiving path 33 of the hot source 32 flows back into theworking chamber 22, increasing the pressure in the working chamber 22.The volume of the energy receiving path 33 is much greater than thevolume of the working chamber 22 in the back-flow phase. Consequentlythe back-flow takes place without any significant pressure drop in theenergy receiving path 33. The temperature of the gas located in the path33 near the outgoing transfer port 27 is at a temperature close to thetemperature of the gas filling the working chamber 22 at the end ofcompression P3. Consequently, the pressure rise during the back-flowphase P4 corresponds essentially to a transfer of mass without any greatthermal impact on the gas.

The back-flow phase P4 is followed by an outgoing transfer phase P5during which the piston 8 t continues its stroke until the workingchamber 22 reaches its minimum volume Vm, whilst the gas present in thechamber 22 is expelled into the cold end of the energy receiving path33. At this time, the pressure in the working chamber 22 increasesslightly, on the one hand to propel the gas, and on the other hand toensure the slight volumetric compression undergone by the gas in thetotal volume of the working chamber 22 and the energy receiving path 33,as this total volume is compressed as a whole by the movement of thepiston 8 t.

When the piston 8 t reaches top dead centre (the closest position to thecylinder head 11), the outgoing transfer port 27 closes and the incomingtransfer port 28 opens, through a corresponding movement of theirrespective valves 29 (FIG. 9B). During this time the intake 26 anddischarge 24 ports remain closed. The pressurised gas originating fromthe hot end of the energy receiving path 33 enters the working chamber22 through the port 28 during an incoming transfer phase P6 until thepiston 8 t reaches a position corresponding to a volume Ve of theworking chamber 22 that is larger than the volume Vs at which theback-flow and outgoing transfer phases P4, P5 started.

The volume Ve is selected so that the mass of gas entering the workingchamber 22 during phase P6 is equal to the mass expelled during theoutgoing transfer phase P5. As the incoming gas is much hotter than theoutgoing gas, this equal mass corresponds to a volume Ve larger than thevolume Vs.

It is important to note that the cycle described and shown in FIG. 4 istheoretical. In practice, the valve controls can be shifted to optimisethe actual cycle in view of the delay for setting gas in motion and theinevitably gradual opening and closing of the valves. For example, inreality, it may be necessary to start the back-flow P4 and outgoingtransfer P5 phase earlier in the cycle to take into account the inertiaof the gases during these movements in one direction and then the other.It is thus possible that in practice the opening point of the outgoingtransfer port, identified by Vs in FIG. 4, moves closer to the closingpoint of the incoming transfer port, identified by Ve in FIG. 4, or evencoincides with Ve, or even precedes it. Indeed, one of the novelfeatures of the invention consists especially of opening the outgoingtransfer orifice whilst the gas contained in the working chamber 22 hasnot yet been compressed to a value corresponding to that of the energyreceiving path 33. Thus, the gas injected into the cold end of that pathis at a remarkably low temperature. This allows for the recovery of moreenergy originating from the exhaust of the internal combustion engine,and even allows, as has been seen above due to the branch 44 (FIG. 1),for the recovery of energy originating from the heat engine discharge.

It is also possible that for a brief instant, the outgoing transfer 27and incoming transfer 28 ports are open at the same time to achieve whatis known as “valve overlap” (well known in four-stroke internalcombustion engines at the end of the exhaust stroke and the start of theintake stroke). The incoming transfer phase P6 is followed by anadiabatic expansion phase P7, during which the piston 8 t moves awayfrom the cylinder head 11 until it reaches its position corresponding tothe maximum volume VM of the working chamber 22. As shown in FIGS. 10Aand 10B, the four ports are closed during this phase and the gasundergoes adiabatic expansion. The diagram in FIG. 4 clearly shows thatthe incoming transfer phase P6 and the expansion phase P7 togethersupply more mechanical energy to the piston 8 t than it consumes tocarry out phases P1, P2, P3, P4 and P5, during which it reduces thevolume of the chamber 22 despite the opposite pressure in the chamber 22during each of these phases. As is known by a person skilled in the art,this excess mechanical energy produced compared with the mechanicalenergy consumed can be seen by the fact that the pressure-volume diagramin FIG. 4 travels in a clockwise direction.

In the theoretical diagram in FIG. 4, it is assumed that the dischargeport 26 opens when the volume of the working chamber 22 reaches itsmaximum value VM (bottom dead centre of the piston 8 t). In practice, asis normal in heat engines with pistons in general, the opening of theexhaust port can occur at an earlier stage to give the pressure time todecrease at least partly before the piston starts its upwards stroketowards the cylinder head.

FIG. 11 shows a four-stroke cycle for the heat engine of the hybridengine according to the invention. In this case, the cycle requires twosuccessive reciprocations of the piston St. One complete stroke of thepiston 8 t towards the cylinder head 11 is given over to exhaust (phaseP1), the next complete stroke is given over to intake (phase P2) untilthe piston reaches bottom dead centre (volume VM). During the intakephase P2, the discharge port is closed. In the example shown in FIG. 11,intake takes place at atmospheric pressure. It is also possible and evenpreferable for intake to take place under pressure by means of aturbocharger such as 37 in FIG. 1, in which case the intake phase P2will be located above the discharge phase P1 in FIG. 11.

The intake phase P2 is followed by a piston St rise phase P23 withoutcompression. This effect can be obtained either by only closing theintake port at the end of phase P23, or conversely closing it before theend of phase P2, so that the end of phase P2 and phase P23 together forma neutralised reciprocation, on either side of bottom dead centre. PhaseP23 is followed by adiabatic compression P3, back-flow P4, outgoingtransfer P5, incoming transfer P6 and expansion P7 phases, which aresubstantially the same as in the two-stroke cycle shown in FIG. 4.

In a manner not shown, the cams 45 associated with the heat engine canbe controlled angularly to adjust the pressure in the energy receivingpath 33. For example, if the pressure in the path 33 falls below a lowerthreshold, the volume Ve at which the incoming transfer port closes isreduced in such a way that less gas can travel from the path 33 into theworking chamber 22.

When the hybrid engine is started after a sufficiently extendedstoppage, the mass of gas present in the path 33 decreases substantiallydue to leaks through the ports 24, 26, 27, 28. The ports are not alwayscompletely tight in the long term, even if the four valves 29 are in theclosed position. The pressure controller in the path 33 automaticallyensures the pressure rise in the path 33 when the hybrid engine isstarted. Even if a pressure controller is not provided, provision can bemade for a pre-defined shifting of the closing point of the incomingtransfer port during an initial phase of operation of the hybrid engineafter each period of stoppage.

FIG. 1 shows two optional features that allow for a reduction in thepressure rise time of the path 33 when the hybrid engine is started.According to a first feature, the two ends of the path 33 can be closedby solenoid valves 46 when the hybrid engine is switched off.

According to an alternative or complementary feature, the path 33 isconnected to a pressure accumulator 47 by means of a solenoid valve 48that on the one hand is controlled to adjust the pressure level in theaccumulator 47 when the hybrid engine has been operating for a certainamount of time (after correct pressurisation of the path 33) and on theother hand allows the accumulator 47 to recharge the path 33 withpressurised working gas (air) rapidly when the pressure in the path 33is insufficient, particularly during the start-up of the hybrid engineafter a period of stoppage.

The maximum temperatures in the heat engine are relatively low, in theregion of 800° C., and can be withstood by appropriate materials,practically without the removal of heat by a cooling system. Thus, toimprove the efficiency of the heat engine according to the invention, itis envisaged that at least some of the surfaces that define the workingchamber 22 will be insulated. To this end, in FIG. 2 the pressure face23 of the piston 8 t is shown to be made up of the external face of aheat insulating coating 49. Such an insulator could also define the bore2 t, at least in its upper part adjacent to the cylinder head 11.

In normal internal combustion engines, heat is discharged equallythrough the cooling system, the lubrication system and the exhaust.According to the invention, it is envisaged that the discharge of heatfrom the internal combustion engine will preferably be through theexhaust, in order to optimise the recovery of energy by the heat engine.

To this end, as shown in FIGS. 12 and 13, provisions are made to reducethe quantity of heat that the exhaust gases from the internal combustionengine transmit to the cylinder head 11 and to its cooling system, notshown. To do this, the cylinder head 11 has a recess 51 the bottom ofwhich is as close as possible to the exhaust port 21 in order to makethe exhaust passage 13, which is formed in the cylinder head 11 andopens out into the bottom of the recess 51, as short as possible. Anexhaust pipe 52 is fixed against the bottom of the recess 51. The innerpassage 53 of the pipe 52 runs on from the passage 13 in the cylinderhead. The pipe 52 is made for example from cast iron or even,preferably, is a double-walled steel passage as shown in FIG. 13.

In FIGS. 14 to 17, the internal combustion engine marked “ICE” andlabelled 54 or 64, and the heat engine marked “HE” and labelled 55 or65, are shown symbolically, together with their respective ports 19, 21,24, 26, 27, 28.

The configuration in FIG. 14 is appropriate for an engine with one heatcylinder and two or three internal combustion cylinders. It is similarto the configuration in FIG. 1, except that the turbine part 56 of theturbocharger 57 is mounted at the output from the energy supply path 34of the hot source 32.

Generally, knowing the quantity of heat that can be recovered in theexhaust of the internal combustion engine, it is proposed in accordancewith the invention that the displacement and supercharging pressure ofthe heat engine be selected so that the mass of working gas processed bythe heat engine during each cycle corresponds substantially to the massnecessary to absorb the heat recoverable from the exhaust of theinternal combustion engine. On a case by case basis, optimisation canconsist of adapting the supercharging pressure of the heat engine,selecting an appropriate number of cylinders, selecting a two-strokecycle or a four-stroke cycle, or selecting a different pistondisplacement for the heat engine from that of the cylinders of theinternal combustion engine.

In the configuration shown in FIG. 15, which is appropriate for a hybridengine comprising a two- or three-cylinder internal combustion engine 64and a three-cylinder heat engine 65, the turbine part 66 of theturbocharger 67 is placed between the discharge port 26 of the heatengine and the branch 44 of the discharge in the energy supply path 34of the hot source 32. This therefore has the advantage of freeing theexhaust of the internal combustion engine, that is, not restricting theexhaust with the presence of a turbocharger. This solution is easier toachieve when the heat engine has multiple cylinders as the pressure inits discharge is pulsed less heavily and thus allows for improvedoperation of the turbocharger.

FIG. 16 shows a configuration that is appropriate for a six-cylinderhybrid engine comprising a three-cylinder internal combustion engine anda three-cylinder heat engine. This configuration will be described withregard to its differences in relation to the configuration in FIG. 14.The turbocharger 57 comprises a compressor part 58 that is supplied notwith atmospheric air but with air that has already been compressed oncein the compressor part 78 of a first turbocharger 77. The turbine part76 of the turbocharger 77 is installed between the exhaust port 21 ofthe internal combustion engine and the catalytic converter 41. Thecharge air produced by the compressor part 78 passes through acharge-air cooler (intercooler) 73 and then reaches a fork 69 at whichsome of the compressed air goes to the compressor part 58 of theturbocharger 57 through a branch 71, and some of the air is sent to theintake 19 of the internal combustion engine 54 through a passage branch72.

The configuration in FIG. 17 is appropriate for a six-cylinder hybridengine comprising a three-cylinder internal combustion engine and athree-cylinder heat engine. This configuration will be described withregard to its differences in relation to the configuration in FIG. 15.The turbocharger 67 comprises a compressor part 58 that is supplied notwith atmospheric air but with air that has already been compressed oncein the compressor part 88 of a first turbocharger 87. The turbine part86 of the turbocharger 87 is installed between the exhaust port of theinternal combustion engine and the catalytic converter 41. The chargeair produced by the compressor part 88 passes through a charge-aircooler (intercooler) 83 and then reaches a fork 79 at which some of thecompressed air goes to the compressor part 68 of the turbocharger 67through a branch 81, and some of the air is sent to the intake 19 of theinternal combustion engine 64 through a passage branch 82.

Thus, in each of the embodiments in FIGS. 16 and 17, the air supplyingthe internal combustion engine is compressed once so that it has aneffective pressure of for example 0.2 MPa (2 bar) whilst the air sent tothe intake 24 of the heat engine is compressed twice to reach aneffective pressure of for example 0.6 MPa (6 bar).

Of course, the invention is not limited to the examples of embodimentsthat have just been described.

The invention relates to all types of heat engine, and is not restrictedto a hot source originating from an internal combustion engine.

In the case of a hybrid engine, the internal combustion engine and theheat engine can have separate engine blocks. The heat engine can beenvisaged as a separate entity that can be fitted to existing internalcombustion engines, or another source of moderate heat.

The invention applies to all types of internal combustion engine, evenengines with a large number of cylinders. In particular, the inventionis particularly advantageous in engines for heavy goods vehicles andships, and also in fixed engines when the heat released by an internalcombustion engine is not recoverable or fully recoverable for uses otherthan the engine.

The efficiency of the hybrid engine according to the invention isgreatly improved compared to the efficiency of a conventional internalcombustion engine. The additional weight and cost and the reduction inspecific power output (power per cubic decimeter of displacement) arecompletely acceptable for most applications, particularly on vehicles,including touring vehicles such as private cars. The industrialinvestment envisaged is limited. For example, a conventional engineblock for an internal combustion engine can be used, in which one ormore cylinders will be allocated to the heat engine.

As the external faces of the valves associated with the transfer ports27 and 28 are subject to the pressure of the path 33, which tends toopen these valves, it can be advantageous to produce them in the form ofpressure-balanced valves, for example according to EP 0 897 059 A2.

The invention claimed is:
 1. A heat engine comprising: an external hotsource having an energy supply path, wherein the engine comprises aworking member and at least one variable volume working chamber for aworking gas, and a distribution mechanism that connects this chamber toa cold input of an energy receiving path during an outgoing transferphase and to a hot output of the energy receiving path during anincoming transfer phase, the energy receiving path cofirgured to heatthe working gas outside the chamber on contact with the external hotsource, wherein the cold input of the energy receiving path is inthermal contact with a cold output of the energy supply path, andwherein the distribution mechanism is configured to: maintain a pressurein the energy receiving path; during stable operation, connect theworking chamber with the cold input of the energy receiving path while apressure in the chamber is lower than a pressure in the energy receivingpath.
 2. The heat engine according to claim 1, further comprising apressure accumulator that is selectively connected to the energyreceiving path of the hot source to accumulate a reserve of pressurisedgas during stable operation and at least partly return this reserve whenthe engine is restarted after a period of stoppage.
 3. The heat engineaccording to claim 1, further including controllable valve timing inparticular to allow for initial pressurisation of the energy receivingpath and/or to control the pressure in the energy receiving path.
 4. Theheat engine according to claim 1, wherein at the start of an outgoingtransfer phase the chamber has a smaller volume than its volume at theend of the incoming transfer phase.
 5. The heat engine according toclaim 1, wherein at least one surface delimiting the working chamber,particularly a pressure face of the piston, is defined by a heatinsulating coating.
 6. The heat engine according to claim 1, wherein theheat engine operates on a two-stroke cycle in which a discharge phase, acompression phase and the outgoing transfer phase follow on from oneanother in a single chamber volume reduction stroke.
 7. The heat engineaccording to claim 6, wherein an intake into the working chamber takesplace under pressure between the discharge phase and the compressionphase.
 8. The heat engine according to claim 1, wherein the heat engineoperates on a four-stroke cycle comprising: a chamber volume growthstroke for a working gas intake phase a chamber volume reduction strokefor the working gas compression phase followed by the outgoing transferphase; a chamber volume growth stroke for an incoming transfer phasethen an expansion phase; and a chamber volume reduction stroke for aworking gas discharge phase.
 9. The heat engine according to claim 1,wherein: the working chamber is delimited by a piston, which alternatelycauses chamber volume growth and reduction strokes; the distributionmechanism is designed to selectively close the working chamber andrespectively selectively connect the working chamber with a working gasintake, a working gas discharge, the cold input of the energy receivingpath and the hot output of the energy receiving path; and in that thedistribution mechanism connects the chamber with the cold input duringthe outgoing transfer phase after the compression phase of the workinggas in the working chamber, and with the hot output during an incomingtransfer phase that precedes an expansion phase.
 10. The heat engineaccording to claim 1, wherein the distribution mechanism is timed sothat at the end of an expansion phase, the chamber has a volume largerthan its volume at the start of a compression phase.
 11. The heat engineaccording to claim 1, wherein a discharge of the heat engine isconnected to a warm input of the external hot source, the temperature ofthe warm input being between the temperature of an energy supply path atthe input of the hot source, and the temperature at which the energysupply path leaves the hot source.
 12. A heat engine according to claim1, wherein the energy supply path is supplied with energy by an exhaustof an internal combustion engine.
 13. The heat engine according to claim12, further including an insulation of the exhaust of the internalcombustion engine over at least part of its path between an exhaust portadjacent to a combustion chamber on the one hand and a hot input of theexhaust gas into the hot source on the other hand.
 14. A hybrid enginecomprising a heat engine according to claim 1 and an internal combustionengine an exhaust of which supplies the energy supply path of the hotsource of the heat engine.
 15. The hybrid engine according to claim 14,wherein the hot source comprises for the energy supply path an outputconnected to the input of a turbocharger turbine.
 16. The hybrid engineaccording to claim 14, further including a turbine is inserted betweenthe discharge of the heat engine and a warm input of the external hotsource, the temperature of the warm input being between the temperatureof the energy supply path at the input of the hot source, and thetemperature at which the energy supply path leaves the hot source. 17.The hybrid engine according to claim 14, wherein the discharge of theheat engine is connected to the input of a turbocharger turbine.
 18. Thehybrid engine according to claim 15, wherein the turbocharger comprisesa compressor the output of which is connected to the intake of the heatengine.
 19. The hybrid engine according to claim 15, wherein theturbocharger is a second turbocharger, the compressor input of which isconnected to the compressor output of a first turbocharger with aturbine through which the energy supply path passes upstream of the hotsource.
 20. The hybrid engine according to claim 19, wherein the turbineoutput of the first turbocharger is connected at least indirectly to thehot input of the hot source.
 21. The hybrid engine according to claim19, wherein the compressor output of the first turbocharger is alsoconnected to the intake of the internal combustion engine.
 22. Thehybrid engine according to claim 14, wherein at least one piston of theheat engine is coupled to a shaft, the internal combustion engine alsocomprising at least one piston coupled to the shaft.
 23. The hybridengine according to claim 14, wherein the internal combustion engine andthe heat engine have a common engine block.
 24. The hybrid engineaccording to claim 22, wherein the at least one piston of the heatengine and the at least one piston of the internal combustion enginehave same pressure face areas and stroke lengths.
 25. The hybrid engineaccording to claim 22, wherein said at least one piston of the heatengine has a pressure face area and a stroke length at least one ofwhich is different from a pressure face area and a stroke length,respectively, of said at least one piston of the internal combustionengine.
 26. The hybrid engine according to claim 14, wherein theinternal combustion engine operates on a four-stroke cycle and the heatengine on a two-stroke cycle.